Compression method and means

ABSTRACT

The invention is a compression method having characteristics of smooth compression and internal cooling of the gas. Embodiments of the invention employ a cylindrical chamber and an orbiting rotor to create a moving duct or chamber whose walls converge, relative to a static gas packet drawn into the moving duct, at a “pinch point”. Preferably the closing speed of the walls is subsonic and the speed of the pinch point is supersonic. This enables high pressure to co-exist, at the narrowing end of the duct, with low pressure elsewhere in the duct, because of the pressure information barrier produced by the supersonic advance of the pinch point. The invention also discloses means for providing gas inlet and outlet functionality, and means for providing a variable flow compressor.

This application relates to the field of gas pumping and compression.

BACKGROUND OF THE INVENTION

Gas compression devices used in refrigeration, air conditioning andindustry consume a large portion of electrical power generated. Anincrease in gas pumping efficiency will result in reduction of carbondioxide emissions. Proposals to sequester carbon dioxide at pressureunderground or in the ocean depths are dependent on using compressionmethods that are efficient and can also overcome problems such as phasechange and the material erosion of compressor parts when compressingimpure gas mixture. Small changes in compressor efficiency may determinecommercial viability.

In an existing compressor design according to international publicationWO 2008/122781 there is provided a compressor comprising a cylinder anda rotor, whereby the rotor traverses the internal circumference of thecylinder and a pinch point is formed at the closest point of the rotorperiphery to the internal wall of the cylinder. The rotor traverses theinternal circumference of the cylinder such that the pinch point movesat high, preferably supersonic speed. In an embodiment, the rotor rollsaround the internal circumference of the cylinder such that the speed ofthe rotor surface, relative to the cylinder wall, is low or zero, thusreducing wear and frictional heating of the components and of the gas tobe compressed, termed herein “rolling”, thus aiding compressorefficiency. Optionally, a strip valve arrangement on the rotor surfaceallows entry of gas into the chamber formed between rotor and cylinder.Optionally, a strip valve arrangement on the cylinder wall allows exitof gas from the chamber and optionally incorporates actuation means tocontrol its opening position.

In another existing compressor design there is provided a compressorcomprising a cylinder and a rotor, whereby the rotor traverses theinternal circumference of the cylinder and a pinch point is formed atthe closest point of the rotor periphery to the internal wall of thecylinder. The rotor moves such that the pinch point moves at high,preferably supersonic speed. The rotor rotates around the internalcircumference of the cylinder such that a fixed point on the rotorperiphery is maintained adjacent to the pinch point—termed herein“rotating”. Optionally, ports in the rotor allow entry and exit of gasvia passages communicating with the axial ends of the cylinder.

Known types of compressors typically suffer from problems which tend toreduce efficiency, including but not limited to those described herein,namely:

-   -   low inter-stage compression rise    -   large physical size relative to gas processing rate    -   Limited ability to offer variable flow rate    -   Complicated inlet/outlet porting and/or valving    -   Losses associated with gas leaks    -   Lowered efficiency related to irregular gas flow    -   Difficulties in achieving balance at high rotational speeds

In particular, efficiency losses can result when a variable flowcompressor is desired. This is at least partly because typical existingvariable flow compressors vary their flow by changing the operatingspeed of the compressor. However, since such compressors are usuallyoptimised for a particular operating speed, changing the operating speedcan result in reduced efficiency.

Variable speed compressors also require expensive control electronics toprovide variable speed drive.

Some existing compressors also suffer from stalling if the inlet chargedensity undergoes a step, or rapid, change.

Other existing compressors achieve variable flow by recirculating orexhausting a portion of the compressed gas, or by throttling the inputof the compressor, however this reduces efficiency.

Many of the above problems also apply to expanders and companders.

BRIEF DESCRIPTION OF INVENTION

The invention is set out in the claims.

In a first embodiment of the invention there is provided a compressor asdefined in claim 1 of the appended claims.

Embodiments of the invention will now be described, by way of example,with reference to the figures which are as follows:

FIG. 1—Schematic view of compressor housing and rotor

FIGS. 2 a to 2 g—‘Rolling’ rotor operation

FIGS. 3 a to 3 g—‘Orbiting’ rotor operation

FIGS. 4 a to 4 g—‘Rotating’ rotor operation

FIG. 5—Strip valve arrangement

FIG. 6—Rotor port arrangement

FIG. 7—Rotor surface features

FIG. 8—Spiral duct embodiments

FIG. 9—Rotor balancing/drive arrangement

FIG. 10—Strip valve and balancing arrangement

FIG. 11—Strip valve actuation arrangement

FIG. 12 a—Dual Lobed Rotor

FIG. 12 b—Dual Lobed Rotor, alternative embodiment

FIG. 13—Disc valve arrangement

FIG. 14 a—Flow control flanges, perspective view

FIG. 14 b-Flow control flanges, axial end view

FIG. 14 c—Angled rotor surface

FIG. 15—Pressure balancing flange

FIG. 16 a—Sliding flange

FIG. 16 b—Sliding flange, alternative embodiment

FIG. 17—Input scoop

FIG. 18 a—Sliding flange, alternative embodiment

FIG. 17—Input scoop

FIG. 18 a—Open channel pump with variable capacity

FIG. 18 b—Closed channel pump with variable capacity

FIG. 19—Outlet redirection feature

FIG. 20—Movable compressor floor, perspective view

FIG. 21—Movable compressor floor, cross sectional view

FIG. 22—Movable compressor floor, cam actuation

DETAILED DESCRIPTION

FIG. 1 shows an existing compressor that has the desired characteristicsof smooth compression and internal cooling of the gas. This compressoremploys a cylindrical chamber (10) and rotor or orbiter (20) to create amoving duct or chamber (40) of unchanging geometry and size, whose wallsconverge relative to a static gas packet drawn into the moving duct(40). In use, (40) walls converge at a lower speed than the point ofclosest approach of the walls [hereinafter called the pinch point (50)]moves along the duct (40). In preferred operation the closing speed ofthe walls is subsonic and the speed of the pinch point (50) issupersonic. As the pinch point (50) advances, the volume in which gas isat highest pressure/temperature also advances to areas of the walls thathave been cooled since last being adjacent to the high temperature gas.When such a compressor is operating with the pinch point (50) moving atsupersonic speeds, information about the pressure rise caused bynarrowing of the duct (40) cannot propagate forward and push the gasforward. This enables high pressure to co-exist, at the narrowing end ofthe duct (40), with low pressure elsewhere in the duct (40) because thevolumes are physically separated by the pinch point (50) and thepressure information barrier (40) produced by the supersonic advance ofthe pinch point (50). This provides a compressor that has the highpressure ratio capability of positive displacement compressors combinedwith the smooth pulse-less outflow of centrifugal and axial machines.

Various embodiments employ a duct (40) created between an innercircumference of a cylinder (10) and a shaped wall (20) moving withinthe cylinder (10) so as to form a narrowing of the duct (40) at thepoint of closest approach of the two members (50).

Three embodiments demonstrating variations on the movement of the rotor(20) within the cylinder (10) will now be described.

As shown in FIGS. 2 a to 2 g, as can be used in a first class ofembodiments described below, a ‘rolling’ rotor (20) rolls around theinner circumference of the cylinder (10) as the rotor (20) traverses theinner circumference of the cylinder (10). The orientation of the rotor(20) is shown by respective arrows A, B, C in FIG. 2 a. The sequence ofsix illustrations shown consecutively in FIGS. 2 b to 2 g illustrates(see arrow A in each) how the orientation of the rotor (20) changes withrespect to the cylinder (10) as the rotor (20) rolls around the innercircumference of the cylinder (10). The rotor changes orientation as itrolls such that the speed of the rotor (20) surface, relative to thesurface of the inner circumference of the cylinder (10) is substantiallylow or zero. The rotor (20) can be arranged to substantially contact theinner surface of the cylinder (10) or the two surfaces can be spacedslightly apart. The rotor (20) can be arranged to roll by means ofcontacting the inner surface of the cylinder (10) or can be rotated byother means such as gears or by entrainment by the gas being compressed.This feature results in a substantially low or zero rubbing speedbetween the surface of the rotor (20) and the inner surface of thecylinder (10), which in turn results in improved wear performance ofthose surfaces. Other results of this feature are lower frictionallosses, lower kinetic energy imparted to the gas being compressed (lowerentrainment) and lower frictional heat imparted to the gas beingcompressed. These results all contribute to greater efficiency of thecompressor.

As shown in FIGS. 3 a to 3 g, as can be used in the first class ofembodiments described below, an orbiting rotor (20) does not changeorientation with respect to the cylinder (10) as the rotor (20)traverses the internal circumference of the cylinder (10). FIG. 3 ashows sequential position 20 a, 20 b, 20 c and correspondingorientations with arrows A, B, C. FIGS. 3 b to 3 g show the sequentialrotor positions and corresponding orientation A. An orbiting rotor (20)results in a greater relative speed between the surface of the rotor(20) and the inner surface of the cylinder (10) than with the rollingrotor (20) of FIG. 2, but a lower relative speed than with a rotatingrotor (20) as will be described in the following paragraph. Efficiencylosses when an orbiting rotor (20) is employed tend therefore to be in arange between those of the rolling rotor (20) and those of the rotatingrotor (20).

As shown in FIGS. 4 a to 4 g, as can be used in the first class ofembodiments or a second class of embodiments described below, therotating rotor (20) changes orientation as the rotor (20) traverses theinternal circumference of the cylinder (10), in such a way that a fixedpoint on the rotor (20) surface A, B, C in the sequential positions 20a, 20 b, 20 c in FIG. 4 a is adjacent to the pinch point (50). Themovement of point A can be seen in the sequential position shown inFIGS. 4 b to 4 g. A rotating rotor (20) results in a greater relativespeed between the surface of the rotor (20) and the inner surface of thecylinder (10) than either the rolling rotor (20) of FIG. 2 or theorbiting rotor (20) of FIG. 3. Efficiency losses when a rotating rotor(20) is employed tend therefore to be in a range which is higher thanthose of the rolling rotor (20) of FIG. 2 or the orbiting rotor (20) ofFIG. 3. An advantage of the rotating rotor (20) of FIG. 4 is that agreater range of valve arrangements can be practically used than withthe other two rotor (20) types. A compressor incorporating the rotatingrotor (20) can be made with fewer moving parts than a compressorincorporating the other two types of rotor.

As shown in FIG. 5, in a first class of embodiments, the duct (40) is achamber formed between two cylinders, one relatively static (10) andacting as a stator and the other (20) acting as a rotor—rolling,orbiting or rotating it within it.

Using a valving mechanism described below, gas is drawn into the duct(40) by a rarefaction caused by the widening of one end the duct (40)(i.e. when the rotor is adjacent an opposing side of the stator). Itpasses through inlets in the walls of either of the cylinders (10, 20)or of the end walls and is expelled at higher pressure at the other endof the duct (40) after it has been compressed by a relative narrowing ofthe duct (40) caused by the orbiting component (20) approaching thestator wall. By mounting blades inside the rotor (20), a degree ofpre-compression can be achieved. In such an embodiment the rotor (20)may have a rolling or rotating surface or may orbit without rotation.

In a device built according to such a first class of embodiments, asshown in FIGS. 5 and 10, there is provided a cylindrical rotor (20),within a cylinder (10). The rotor (20) is provided with a surfacechannel (210), of depth equal to the thickness of strip (220) that fitswithin the channel (210). The strip (220) is of larger circumferencethan the rotor (20) circumference, so that when the strip (220) ispressed onto the rotor (20) it forms a gas tight seal. However becausethe strip (220) is of larger circumference than the rotor portion (20),the strip (220) will always protrude above the rotor (20) surfacecircumference, away from the final point by virtue of the squeezingforce exerted on it there, allowing gas to flow through openings (230)in the base of the channel (210) to outside the rotor (20).

The cylinder (10) is similarly provided with a channel (240) and strip(250) on the outside, allowing gas to pass from the inside of thecylinder (10) to ducting means on the outside. This outlet strip (250)may be provided with reinforcement across its width to support itagainst high gas pressures.

In operation the rotor (20) orbits, preferably at a speed that resultsin the pinch point (50) between the rotor (20) and cylinder (10)rotating at supersonic speed.

As the pinch point (50) rotates, low pressure follows behind (in termsof the direction of rotation) the pinch point (50), pulling the stripvalve (220) away from the rotor (20) and continuously inducing gas intothe chamber (40). At the other end of the chamber (40) the convergingsurfaces of rotor (20) and cylinder (10) compress previously inductedgas and force it out of the chamber (40) through the exit strip (250),which is forced to and held in an open position by the pressure of gasin front of the pinch point (50). To prevent the exit strip (250)overlapping the pinch point (50) and allowing gas to escape from thehigh pressure volume into the low pressure volume, the exit strip (250)may be actuated by mechanical, electrical or magnetic means to controlthe distance of its opening (270) from the pinch point (50). As shown inFIG. 11, a cam (261) on the drive shaft (660) operates a pushrod (260)which operates to lift the exit strip (250). This actuation is alsohelpful for controlling start and shutdown conditions and to give adegree of capacity control. In rolling or rotating rotor embodiments(see below) where the high pressure exit side of the pinch point (50)can be separated by some distance from the low pressure inlet side,actuation of the strip (250) may be used to restrict the area of theoutlet by moving the opening (270) partially past the pinch point (50)and so controlling the pressure ratio of the device. Hence, in anembodiment, a strip is deformable by mechanical actuation, in particularby an actuator such as a cam and pushrod coupled to the rotor, forexample the rotor drive shaft.

As shown in FIG. 6, a blind passage (275) or passages are providedwithin the rotor, open on the axial face and terminating adjacent theinside of the rotor surface. This passage (275) communicates with theaxial face of the rotor (20), so that cooling fluid may be circulatedbehind the rotor (20) circumferential surface. The walls and end platesof the chamber (10) are additionally provided with passages for thecirculation of cooling fluid. Finned means (276) may be provided toincrease the heat flow from the chamber (10) walls to be cooled into thecooling fluid.

In operation the rotor (20) is rotated with the inlet conduit (330)leading, so that the duct (40) rotates with the rotor (20) at a speedpreferably in excess of the local speed of sound. Appropriate curvatureof the inlet conduit (330) passage way causes gas to be drawn from anaxial face of the rotor (20) into the conduit (330) in a substantiallyradial direction. As the duct (40)—i.e. the space between rotor andstator—rotates around the stator, the gas is confined to a convergingduct (40) formed between the surface of the rotor (20) and the cylinder(10) wall. The supersonic speed of the approaching pinch point (50) doesnot give time for information about increasing pressure to propagateupstream. The gas is steadily compressed until, as the pinch point (50)reaches the gas, it is permitted to escape at high pressure through theoutlet (340) and through passages within the rotor (20) to a radial endof the rotor (20) from whence it is ducted out of the device. Duringcompression in the duct (40) the gas temperature increases. The heat ofcompression is transferred continuously, both through the wall of therotor (20) into the cooling fluid circulating behind the wall andthrough the chamber walls (10) into the cooling fluid (277) circulatingthere.

As shown in FIG. 7 the surface of the rotor (20) may be provided withspiral grooves (400) and/or passages (410) to conduct high pressure gasthat passes the pinch point (50) or along the axial ends of the chamberback to a selected or controlled point (420) in the duct (40). This gasis cooled on its passage back to the chamber and this is moreadvantageous for the efficiency of the device than allowing it tore-emerge at the inlet end of the chamber (330). In complex devices itwould be possible to bleed this gas through micro pores (430) in therotor surface to promote laminar flow.

In this second class of embodiments the device may include a rotor (20)with the converging duct (40) formed between the cylinder (10) wall andthe surface of the rotor (20) (as shown in FIG. 6) or the convergingduct (40) may be formed between the cylinder wall (10) and a channel ofreducing cross-section (330) on the rotor (20) where the rotor isconcentric with the stator (as shown in FIG. 8).

Referring, for example, to FIGS. 7 and 8, in order to avoid acceleratinggas, it is important that for a given cross-section of duct (40) theratio of stationary to moving duct (40) surfaces should be as high aspossible. In embodiments using a channel (330) within the rotor (20),the duct (40) may be formed by a groove (580) which winds spirally downthe circumferential face of the rotor (20) so that all parts of the duct(40) including the high pressure/high temperature end of the duct (40)are continuously exposed to fresh surface areas to conduct heat from theduct (40). In such an embodiment heat transfer up the cylinderstationary wall (10) may be reduced by flanges (350) behind the surface(see FIG. 6). The rotor (20) may be further cooled by internal fluidflow along the sides of the duct (40) and side of rotor (20).

As shown in FIG. 8, a second class of embodiments employ a rotatingrotor (20). In such a device there is provided a rotor (20) within acylinder (10), the rotor (20) being profiled so that a substantial partof the rotor (20) circumferential surface remains in rotatably closeproximity to the inner wall of the cylinder (10) as it traverses theinner wall. The remaining circumferential surface of the rotor (20) isshaped or cut out so as to create a duct or groove (40) with a narrowingend (530) between it and the cylinder (10) wall. A wider end of the duct(540) is provided with an inlet conduit (520) communicating with thecentral part of an axial face of the rotor (20). Spaced from the duct(40) the circumferential surface of the rotor (20) is provided with anexit conduit (550) communicating with another portion of the axial faceof the rotor (20). In large devices there may be provided more than oneshaped duct (40).

In a spiral duct (500) embodiment the output pressure ratio may becontrolled by providing a moveable sleeve (510) between the rotor (20)and cylinder (10). In operation, gas inlet (520) is through one axialend of the chamber (40) and outlet (550) through the other. Moving thesleeve (510) axially with respect to the rotor (20) changes the outletarea and so changes and controls the pressure ratio of the device.

Any of the above embodiments may be provided with means to adjust theoffset of the rotor (20) from the central axis of the containingcylinder (10) and so adjust the clearance between the rotor (20) andcylinder (10) at the pinch point (50). This is advantageous for wearcompensation, adjusting for different rates of thermal expansion,reducing leakage and to control capacity.

As shown in FIG. 9, an arrangement for driving the assembly, andadditionally adjusting the offset of the rotor (20) from the centralaxis of the containing cylinder (10), and thereby adjusting theclearance between the rotor (20) and the cylinder (10) at the pinchpoint (50), is described herein. In overview, the rotor (20) has a rotoraxis (670) each end of which is coupled to a drive rotor support and anidler rotor support (680, 690) respectively, each of the drive rotorsupport and the idler rotor support (680,690) in turn are coupled to adrive shaft and an idler shaft (660, 650) respectively, which arearranged such that they are on the central axis of the cylinder (10) andare each supported by a bearing support (630).

In more detail, an end of the rotor axis (670) is joined by a coupling(600) to a drive rotor support (680), and an other end of the rotor axis(670) is joined by a coupling (600) to an idler rotor support (690). Theidler rotor support (690) is joined by a coupling (600) to a fixed shaft(650). The drive rotor support (680) is joined by a coupling (600) to adrive shaft (660). Both drive shaft (660) and idler shaft (650) arearranged to be parallel to the rotor axis (670) and to lie on thecentral axis of the cylinder (10). Each rotor support (680, 690) isarranged to support the rotor axis (670) such that the rotor (20)surface is substantially positioned close to the inner circumference ofthe cylinder (10). Both idler shaft (650) and drive shaft (660) aresupported by a bearing support (630) and are rotatable within, andaxially constrained relative to said bearing support (630). Each bearingsupport (630) is arranged such that its axial distance from the centreof the rotor axis (670) is equal to that of the other bearing support(630) and is controllable. By controlling the distance of the bearingsupports from the centre of the rotor axis (670) it is possible to varythe position and angle of each rotor support (680, 690) and resultantlyit is possible to vary the running clearance between the rotor (20) andthe housing (10).

Three classes of coupling (600) can be advantageously employed in thepreceding arrangement. A first class of coupling (600) includescouplings which are suitable for forming a joint which is articulated intwo axes between two shafts, but not capable of transmitting any axialtorque. An example of a commonly known coupling (600) falling into thefirst class is a ball joint. A second class of coupling (600) includescouplings which are suitable for forming a joint which is articulated intwo axes between two shafts, and capable of transmitting axial torque.An example of a commonly known coupling (600) falling into the secondclass is a constant velocity joint, a Hardy-Spicer universal joint,certain types of rubber couplings or compliant rubber tubing. A thirdclass of coupling includes couplings witch are suitable for forming ajoint which is capable of articulating in one axis and capable oftransmitting axial torque. An example of such a joint is a hinged joint.

The drive shaft (660) can transmit rotational torque via a drivecoupling (640). The drive shaft (660) is coupled to the drive rotorsupport (680) by a coupling (600) of the third class. The end of thedrive rotor support (680) which is coupled to the rotor axis (670) isthereby constrained to orbit in a circular motion around the draft shaft(660) axis.

In embodiments employing a rolling rotor, the drive rotor support (680)is coupled to the rotor axis (670) by a coupling (600) of the firstclass. The rotor axis (670) is coupled to the idler rotor support (690)by a coupling of the first or second class. In such embodiments, eitherat least one of the coupling (600) which couples the rotor axis (670) tothe idler rotor support (690) and the coupling (600) which couples theidler rotor support (690) to the idler shaft (650) are of the firstclass, and/or the idler shaft (650) is free to rotate. The rotor (20) isthereby free to roll independently of the drive shaft (640) orientationand the idler shaft (650) orientation, but the rotor (20) is compelledto traverse the inner circumference of the cylinder (10) by the drivetransmitted from the drive shaft (640) to the drive rotor support (680).

In embodiments employing an orbiting rotor, the drive rotor support(680) is coupled to the rotor axis (670) by a coupling (600) of thefirst class. In such embodiments, both of the coupling (600) whichcouples the rotor axis (670) to the idler rotor support (690) and thecoupling (600) which couples the idler rotor support (690) to the idlershaft (650) are of the second class, and the idler shaft (650) is fixedso that it cannot rotate. The rotor (20) is thereby constrained so as tomaintain its orientation with respect to the cylinder (10) by virtue ofits connection to the fixed idler shaft (650). The rotor (20) iscompelled to traverse the inner circumference of the cylinder (10) bythe drive transmitted from the drive shaft (640) to the drive rotorsupport (680).

In embodiments employing a rotating rotor, the drive rotor support (680)is coupled to the rotor axis (670) by a coupling (600) of the second orthird class. The rotor axis (670) is coupled to the idler rotor support(690) by a coupling of the first or second class. In such embodiments,either at least one of the coupling (600) which couples the rotor axis(670) to the idler rotor support (690) and the coupling (600) whichcouples the idler rotor support (690) to the idler shaft (650) are ofthe first class, and/or the idler shaft (650) is free to rotate. Therotor (20) is thereby constrained to maintain its orientation withrespect to the drive rotor support (680), and is unconstrained relativeto the idler rotor support (690) orientation, and as a result, a fixedpoint on the rotor (20) surface is maintained adjacent to the pinchpoint (50). The rotor (20) is compelled to traverse the innercircumference of the cylinder (10) by the drive transmitted from thedrive shaft (640) to the drive rotor support (680).

Although the rolling, orbiting and rotating rotor constraintarrangements have been herein described with reference to the use ofspecific combinations of the aforementioned classes of coupling, it willbe appreciated that the rotor characteristics described herein can beaccomplished by other combinations not described. Accordingly, thedescriptions of the orbiting, fixed, and rotating rotor constraintarrangements described herein are not intended to be limiting to thescope of the invention, the invention being set out in the claims.

As shown in FIG. 9, a means for counterbalancing the rotor (20) isprovided. The drive rotor support (680) is extended past the coupling(600) which couples the drive rotor support (680) to the drive shaft(660), in a direction away from the rotor (20). A counterbalance weight(620) is provided either separately, or integrally with the drive rotorsupport (680) extension. Similarly, the idler rotor support (690) isextended past the coupling (600) which couples the idler rotor support(690) to the idler shaft (650), in a direction away from the rotor (20).A counterbalance weight (620) is provided either separately, orintegrally with the idler rotor support (690) extension. Eachcounterbalance weight (620) is arranged to have a weight and a distancefrom the central axis of the cylinder (10) such that the weight of therotating components on the opposite side of the central axis of thecylinder (10) is balanced. The mass or position of the counterbalanceweights (620) can be adjusted during operation of the compressor, tocompensate for thermal expansion or other effects which would otherwiseupset the balance of the rotating components of the compressor. This canbe achieved by the use of actuators to adjust the position of thecounterbalance weights (620) on the rotor supports (680, 690),Alternatively, the mass of the counterbalance weights (620) can bealtered, for example by pumping fluid or gas in or out of thecounterbalance weights (620) which can incorporate a fluid or gasreservoir.

FIG. 10 shows an alternative arrangement for counterbalancing the rotor(20) where the drive shaft (660), rotor axis (670) and counterbalanceweights (620) are housed within the cylinder (10), this beingadvantageous in that sealing of the chamber is facilitated.

Although the manner in which the various chambers are sealed and ductedare not described in all cases in detail it will be appreciated that inembodiments of this invention the usual sliding seal means of thecompressor art are provided to prevent leakage of gas from high pressurevolumes to low pressure volumes. Ducting means to direct low pressuregas into devices and high pressure gas away from the device are alsoprovided.

In any of the above embodiments conventional control means of the art,such as valves, may be used in combination to control and regulate flow.

Although embodiments have been described with a static cylinder (10) anda movable rotor (20), other embodiments may employ a moving cylinder(10) and static rotor (20) or both moving rotor (20) and cylinder (10).

A compressor may be reversed, with appropriate valving, to operate as anexpander.

Advantages of such a compressor as described above, and otherembodiments, are that high efficiency of compression and high stagepressure rise are achieved by compressing gas while imparting as littlekinetic and friction energy to the gas. Embodiments also allow coolingof the gas while it is being compressed.

In axial and centrifugal compressors the necessity of multiple stages,caused by the low pressure rise per stage can be exploited to provideinter-cooling between stages. For high efficiency of compression allsurfaces enclosing the gas may be cooled and the gas and/or surfacescontinuously changed so that the gas is brought into contact withfreshly cooled surfaces during compression. Preferably the gas shouldnot flow relative to the walls as this causes frictional heating.

Further advantages of embodiments follow:

By employing supersonic rotation of the pinch point, the simplemechanical layout of the invention is made possible, since high pressurecannot propagate to the low pressure areas of the chamber and thereforeno mechanical separation between low and high pressure regions of thechamber is required.

The continuous rotational compression means of the invention allows forsmooth continuous compression. By employing smooth and continuouscompression means, the invention advantageously reduces the energyimparted to the gas being compressed.

By employing adjustable running clearance means, and/or a rotor whichrolls as it traverses the internal circumference of the cylinder,frictional losses are reduced, which reduces heating of the gas to becompressed and thereby increases efficiency.

The fixed chamber volume of some embodiments allows for enhanced heattransfer properties because the maximum chamber surface area is alwaysin contact with the gas being compressed. This allows the gas beingcompressed to be more effectively cooled, which in turn aids compressorefficiency.

The amount of gas processed in each revolution is greater than thevolume of the interior volume of the cylindrical chamber and the volumeof the rotor. The swept volume is the cylinder volume less the volume ofa rotor having a radius equal to {radius of the rotor minus the radialoffset of the rotor axis from the cylinder axis}. In other words, thesweep path of the rotor surface diametrically opposite the pinch pointdefines the swept volume.

A further advantage of a compressor according to embodiments is that itexhibits high flow properties compared to, for example, axial orcentrifugal compressors of a similar physical size. As a result, acompressor can be made physically smaller than some other knowncompressors.

Further developments in compressor technology will now be described indetail below.

A dual-lobed rotor embodiment of the compressor described above is shownin FIG. 12 a. An elliptical rotor 20 within a circular housing 10, orinner cylinder, has an axis of rotation. The clearance between theelliptical rotor and the housing forms a pair of ducts 40, each boundedby a pair of pinch points 50 at the parts of the rotor 20 having thefurthest radius from the axis. The rotor 20, in embodiments,incorporates a pair of inlet flow passages 1200 arranged to communicateaxially with the housing and with the duct 40 via the surface of therotor 20. In use, flow passage 1200 carries inlet fluid (e.g. gas) froman axial end of the compressor and discharges the fluid into the inletend of the duct 40. The gas is then compressed in the way previouslydescribed for the single duct embodiments above. An advantage of thedual lobed embodiment is enhanced balancing properties at highrotational speeds.

Alternatively, the rotor 20 may be shaped as shown in FIG. 12 b, so asto provide a different profile of duct cross-sectional area variationwith respect to angular position of the rotor, the rotor having acomparatively narrow region near the axis and lobes at each radial end.In use, the inlet flow passages 1200 (not shown in FIG. 12 b butequivalent to those of FIG. 12 a) advantageously impart a degree ofcentrifugal compression to the inlet gas and aid inlet charge filling ofthe compressor duct(s) 40, thus allowing a reduction in rotational speedfor a given required compression capability. The ends of the rotor lobesare optionally flattened so as to enhance separation of the highpressure gas close to the pinch point 50 in front of the advancing pinchpoint 50 from the low pressure gas behind the moving pinch point 50.This helps to reduce leakage past the pinch point 50. In a furtherdevelopment, as shown in FIG. 13, high pressure gas in front of theadvancing pinch point 50 is allowed to leave the compressor by theprovision of resilient deformable discs 1300 in the axial end of thehousing 10, so as in use to seal against the axial ends of the housing10 under low or negative internal pressure, but arranged to deformoutwardly and allow high pressure gas to exit the housing 10. Thisprovides a simplified output valving arrangement.

The above embodiments comprise a compressor having a rotor 20 within astator chamber (or housing) 10, where the rotor 20 is shaped so asclosely approach or touch the chamber wall at one or more pinch points50, the rotor 20 having no abrupt changes of surface direction betweenpinch points 50. When the rotor 20 operates at supersonic speed highcompression ratios can be achieved. Variations include operating thecompressor with both a rotor surface that rotates relative to the statorchamber 10 and a circular surface that rolls relative to the statorchamber 10 wall. The latter is the most efficient mode of operation butmore complex to manufacture.

An improvement, according to the present invention, which simplifies theport arrangement in these devices and to reduces the reduction ofleakage from high pressure volumes to low pressure volumes will now bedescribed. Embodiments also relate to providing an unimpeded gas paththrough the device and to reduction of leakage.

Reduction of leakage has previously involved very tight tolerancesbetween high speed surfaces. If they touch, severe damage may result.

In a simple compressor shown in FIG. 14 a, the rotor 20 has one or morepinch points 50 with no abrupt changes of surface between pinch points50 and rotates within a stator chamber 10. The portion of the rotor 20which extends relatively close to the stator 10 wall (thereby formingthe pinch point 50) may be wide, i.e. the pinch point 50 need not be anarrow region of close approach of the rotor 20 to the stator wall 10,but can optionally be a relatively wide area on the rotor 20circumference. This pinch point portion 50 may have axially orientatedgrooves 1400 to create a labyrinth seal against fluid flow in acircumferential direction. The rotor is provided with rotatable sideflanges 1410, 1420 (inlet=1410, outlet side=1420), on both axial sidesof the rotor 20, that extend radially to closely fit within the chamber10, and between the rotor portion 20 is eccentrically mounted. Theperiphery of the flanges 1410, 1420 are provided with a series ofcircumferential grooves 1430 to provide labyrinth seals against fluidflow in an axial direction. The stator housing 10 is also optionallyshaped to improve sealing.

A small section 1440 of the flange 1410, 1420 on one axial end is cutaway near the leading side in the direction of rotation of a pinch point50. This provides an outlet (or exhaust) port 1440. A larger section1450 of the flange is cut out or reduced in diameter on the other axialend of the rotor 20, so as to form a gap, near the trailing side in thedirection of rotation of a pinch point 50, to provide an inlet port1450. This may extend to where the rotor 20 surface is furthest from thestator chamber wall 10 (point 1480 in FIG. 14 b), i.e. almost 180degrees around the rotor 20 circumference (for a single pinch point 50embodiment). The cut out outlet 1440 and inlet 1450 should be designedto be aerodynamically efficient and may also be shaped to encourage flowonto and off the rotor 20 surface. The outlet cutout 1440 is ideallysituated at a high pressure region close to the pinch point 50. Theinlet cutout 1450 is ideally situated away from the high pressure regionso as to reduce shock waves reflected into the inlet port of thecompressor (not shown). The rotor 20 surface may also be angled oneither side of the pinch point 50 to reduce dead flow volume (i.e.portions of the volume of the duct 40 which are aerodynamicallyrelatively inaccessible to fluid due to their location in sharplydefined recesses) as shown in FIG. 14 c.

In order to prevent flow in the circumferential labyrinth grooves on theflanges providing a leak path into the cut away section, the labyrinthgrooves are sealed at or near the edge of the cut away sections (asshown in FIG. 14 a, item 1405). To prevent damage from touch downbetween the tips of the seals on the flanges and the chamber wall and toallow the compressor to run-in to a very gas tight condition, abradablematerial can be used in sections of the chamber wall 10 adjacent therotating flange labyrinths 1430 of the rotor 20.

When run at high pressure ratios, considerable thrust is produced at thehigh pressure end of the compressor. This may be counteracted byproviding a second sealed flange wall 1540 in the outlet chamber 1530,as shown in FIG. 15.

In operation the rotor 20 turns at high speed, gas flow is inducedthrough the larger inlet cut away 1450 by a rarefaction following thepinch point 50. Induction ceases where the inlet 1450 ends and therotor/stator gap is largest (1480 in FIG. 14 b). Inducted gas is now ina converging duct 1490 and is compressed and forced out of thecompressor in front of the approaching pinch point 50, via the outlet1440. Because a high relative speed of gas and rotor 20 is desirable,the gas should enter the compressor on a path that is static relative tothe stator 10 or angled to oppose the rotor 20. Gas flows in an almoststraight path through the compressor, relative to the stator 10, andmoves relative to the rotor 20.

A further improvement will now be described which provides for changingthe flow of compressors and pumps without altering the rotor speed. Thisis advantageous as it removes the need for complicated and expensivevariable speed drive control apparatus. In existing positivedisplacement compressors this is commonly achieved by throttling theinflow or re-circulating flow. These methods reduce the efficiency ofsuch a device, and in order to provide an improved device the presentinvention alters the displacement by changing the cross-sectional areaof the flow channel or duct 40 while the compressor is operating. Thisalso changes the swept volume per rotation. Thus the flow per revolutionof the rotor 20 can be adjusted to anywhere between the maximum designflow and zero flow.

Adjustment of displacement in devices with open gas ports disclosedabove may be achieved by axially moving one or both of the flanges 1410,1420 so that the cross sectional area of the flow channel (or duct) 40is changed.

In a simple embodiment of this invention as shown in FIG. 16 a there isprovided a flange 1410 at the inlet axial end of the rotor 20. Thisflange may be an integral part of the rotor 20 or fixed to the rotor 20.A second flange 1620 is fixedly mounted on a first end of a hollowmember 1610 that fits closely round the rotor 20 so that the hollowmember 1610 can slide inwards and outwards on the rotor 20 to vary theseparation of the flanges 1410, 1620. In operation the rotor 20 turns,gas is inducted into one end of the chamber 1640 and forced out at theother as described above. To vary the displacement and so change thevolume of fluid that is pumped per rotation, the hollow member 1610 andthe flange 1620 are moved towards or away from the inlet flange 1410 bymoving the bearing housing 1660 at the output end of the device andcausing the hollow member 1610 to slide along the drive shaft 1280 towhich the rotor 20 is fixed. A splined connection between the hollowmember 1610 and the shaft 1280 prevents torque from the drive beingtransmitted entirely through the rotor 20 to the hollow member 1610 anddistorting it and unbalancing the assembly. As will be apparent to theskilled reader, movement of the bearing 1660 can be achieved by anyappropriate mechanical, pneumatic, hydraulic or magnetic means. Ofcourse, the inlet flange 1410 could be moved instead of or as well asthe outlet flange 1620.

An alternative arrangement is shown in FIG. 16 b, in which the movableflange 1620 is supported on the rotor by means of opposed engaging ends1630, 1640 axially movable in cooperating axial slots 1650, 1660 in therotor 20, thereby obviating the need for a hollow member 1610 on whichthe flange 1620 is supported in the embodiment shown in FIG. 16 a. Theflange 1620 is shaped to fit closely between the rotor 20 and thehousing 10, with an outside shape similar to that of the inside shape ofthe housing 10, and an inside shape similar to that of the surface ofthe rotor 20, thereby to minimise leakage past the flange 1620.

As shown in FIG. 17, flow into the device may be encouraged by anglingthe leading edge of the inlet opening 1450 of the flange 1410 on theinlet side, away from the rotor 20 in an axial direction, so as to forma scoop for encouraging fluid into the chamber (or duct) 1640. Thisangle may be varied during operation by appropriate mechanical orelectromechanical actuation means to suit the desired flow.

Clearly a device of this construction can also be used at moderatespeeds for pumping liquids. Variable displacement at a fixed speed isuseful in applications where variable proportions of fluids have to bemixed in industrial processes.

Fluid flow through a device as described can also be reversed so thatthe device functions as a variable expander or turbine to extract workfrom fluid flowing from high pressure to low pressure.

Furthermore connecting a variable displacement compressor according tothis invention with a variable displacement expander according to thisinvention enables work to be recovered from the expansion process (forexample, a refrigerant expansion process) rather than energy beingwasted in throttled expansion (i.e. operation in expansion mode with athrottled, or partially closed off, inlet port). This recovered work canemployed to reduce the work required for compression by a direct driveconnection to the compressor. This is not normally practicable incompression systems because of the varying volumes that have to beprocessed by compressor and expander. However, employing independentlyvariable compressor and expander displacements overcomes this problem.Thus higher efficiency can be achieved in refrigeration processes withunprecedented control over conditions in the condenser and evaporator.Such control is also required in Fuel Cell systems where the volumetricflow in and out of the compressor varies so significantly that it hasnot hitherto proved practical to directly feed the work recovered in theexpander to the compressor.

Further, in existing pumps and fans it has been known to change flow bydynamically changing the angle of attack of rotating blades and bychanging the angle of non rotating vanes at the inlet to the rotor, soas to change the angle at which flow impinges on the turbine blades.These methods have the problem that they reduce efficiency by imposingan unnecessary acceleration on the fluid. The present method of alteringthe channel hydraulic cross-sectional area described above can also beapplied to turbines both with enclosed or shrouded flow channels andopen or unshrouded flow channels.

In another embodiment of a device where displacement is adjusted byvarying the hydraulic cross-section of a flow channel as shown in FIG.18 a, as applied to unshrouded and shrouded turbine flow channels of atype commonly used in liquid pumps, there is provided a shaft 1280 onwhich is mounted a rotor member 1810 with channels 1825 formed betweenstrakes 1820. Slideably mounted on the shaft is a second rotor member1830 with projecting members 1840 that fit closely within the channelsof rotor member 1810. In operation, axially moving the second rotormember 1830 that has projecting members 1840 moves the projectingmembers 1840 in and out of the channels between the strakes 1820 tochange the hydraulic area of the channels encountered by flow G. In openturbines the projections form the floor 1850 of the open channel and aremoved axially in and out to change the effective hydrauliccross-sectional area formed by the rotating channel and the statorchamber wall. The projection forming the floor may have a curved surfaceto more approximately conform to a circular cross-section and somaximise the flow area to wall area and so reduce friction losses. Analternative embodiment having closed channels is shown in FIG. 18 b.

In devices built according to this invention it is not always necessaryto actuate the sliding member in the direction that widens the channelas there is sufficient pressure in the channel to provide a returningforce to widen the channel again. Actuation is only required to narrowthe channel and this will normally be achieved by mechanically pushingon the moveable member 1830 from an axial direction (for example bydirect, hydraulic or pneumatic action) or by advancing a magnetic fieldto repel the moveable member 1830). Of course, movement in eitherdirection can be augmented or opposed by appropriate biasing means suchas a spring.

Compressors and pumps can generally operate in reverse to act as enginesby converting pressure and kinetic energy into work. The abovedescription of varying flow by changing the width of the channel on therotor 20 is also industrially useful in such engines.

Moving the bounding surfaces of a rectangular cross sectional duct 40 ina compressor or pump changes the cross-sectional area and so can beemployed to change the fluid flow without changing the rotational speedof the device. This offers advantages over changing fluid flow bychanging rotational speed, or by throttling the inlet, which are theestablished methods of providing variable flow. This method of varyingthe volume displaced per revolution can be applied to converging ordiverging ducts within a compressor or pump, as well a ducts ofunvarying cross-sectional area.

In a compressor intended to operate at supersonic speed of rotation, thehigh speed stresses in the material are challenging. This isparticularly the case when moving the radial plane boundaries (i.e.flanges 1620 defining a duct 40 boundary) in an axial direction. Aproblem with moving flanges 1620 axially is that axial movement alsoalters the axial balance of the rotor and induces vibration. A furtherembodiment has been developed in which the inner boundary of the duct40, comprising the rotor 20 surface, hereinafter termed the floor 2000,is moved to adjust the flow through a duct 40. This overcomes the stresschallenges of moving the radial plane boundaries and also enablesstreamlining of the inlet to reduce entry turbulence losses.

In an embodiment of such a device (FIG. 20 & FIG. 21, showing one ductonly) adapted for compressing gas, there is provided a rotor 20 with twohalf crescent ducts 40. The or each duct 40 is bounded at the outersurface by the circumferential wall of the cylindrical chamber 10, inwhich it rotates. The rotor 20 is provided with two side walls (orflanges 1410, 1420) extending in the radial plane which define the axialwidth of the duct 40. Either of these walls (or flanges 1410, 1420) maybe tapered inwards towards the inlet 1450 and outlet gas ports 1440which are provided at each end of the duct 40. This improves gas flow.An inlet port 1450 is provided through a first side wall 1410 at thelarger cross-sectional end 2010 of each duct 40 and communicates with asource of gas to be compressed. An outlet port 1440 is provided in thesecond side wall 1420 at the smaller cross-sectional end 2020 of eachduct. The outlet port 1440 communicates with a collection and diffusionchamber 1530 which is bounded by the second side wall 1420 and abalancing side wall 1540 which acts to cancel the axial thrust otherwisecreated by high pressure gas. High pressure gas exits radially throughoutlet passages 2150 arranged in the stator housing 10 adjacent to thediffusion chamber 1530.

The floor 2050 of the duct 40 is shaped to provide a converging duct 40towards the outlet 1440, in co-operation with the chamber wall 10. Thefloor 2050 is supported and restrained at the outlet end of the duct2020 by means such as a hinge or pivot axis 2060 which defines an axiswhich is near the outer circumference of the rotor 20 and about whichaxis the floor 2050 can rotate towards or away from the rotor centrethrough a small arc. The inlet end 1450 of the duct 40 is bounded by anangled or tapered surface 2070, that is arranged so as to smooth thepath of inlet fluid into the duct 40, for example, a duct whichgradually widens in a rotor axial direction towards the area near theinlet port 1450.

The side wall 1410 on the inlet side is provided with an opening 2080 tocreate an inlet port 1450. In operation the floor 2050 is moved aroundits pivot axis 2060 to change flow by changing the cross-sectional areaof the duct 40, while leaving the overall geometry of the chamber 40substantially unchanged. The opening 2080 must be large enough to act asan inlet port 1450 of suitable size for the designed flow. To smoothflow in the entry chamber of the compressor/pump and into the port, aplate 2090 of the same width as the opening 2080 may be attached to theaxial end of the floor 2050 or be integral to it, so that, as the floor2050 moves to and fro, the plate 2090 closes the un-opened area levelwith the outer surface of the side wall 1410 and so reduces turbulencein the entry chamber adjacent the inlet side wall 1410.

The entry chamber (not shown) may also be provided, as is known in theart of pumps and compressors, with angled inlet vanes or ducts either onthe stator chamber 10 or rotating as part of the rotor 20, or both, toturn the inflow in a direction that increases efficiency (i.e. contraryto the direction of motion of the pinch point 50). In the device theintention is to impart velocity to the inlet gas in the oppositedirection to the rotor direction, so as to lower the speed that therotor 20 has to turn to create super sonic speed relative to the gas.Such pre-rotation devices may also include angling the trailing edge ofthe inlet 1450 outwards to induce gas into the duct as shown in FIG. 17.

Similarly the outlet port 1440 may be provided with angled vanes tocompound some of the kinetic energy of the fluid leaving the chamber 40,as shown in FIG. 19. In the embodiment described here the fluid (G) willleave the duct (chamber 40) with high kinetic energy in an axialdirection and will be slowed in the diffusion chamber 1530. In somedesigns it may be advantageous to turn this flow using a vane 1910 (orramp) or vanes against the direction of rotation to provide a thrust (F)on the rotor 20 in the direction of rotation, as shown in FIG. 19. Thisacts both to slow the fluid (G) and increase its pressure and to recoverenergy from the fluid. Advantageously, in a supersonic compressor,wherein the pinch point 50 is moving at or in excess of supersonicvelocity, such energy recovery is possible without imparting backpressure to the inlet.

In the embodiments of FIGS. 20 and 21, the outer edges of the duct 40walls may be angled inwards to prevent the floor 2050 moving too faroutwards and touching the chamber wall 10.

In operation it is preferable to have low pressure beneath (i.e.radially inwards of) the floor 2050 because this enables higher pressurein the outlet end of the duct 2020 to act on the floor 2050, pushing itinwards to counteract centrifugal forces. To achieve this, the floor2050 may be sealed against passage of gas from the duct 40 to the innerpart of the rotor 20. The seals may take the form of lip seals (notshown) mounted on sides of the floor 2050. The lips are pressed againstthe rotor 20 side walls (4, 4 a) by pressure in the duct 40. To adjustthe level of the floor 2050 without requiring high force to overcome thefriction of the seals, the lip seals may be released by admitting highpressure gas into the interior of the rotor 20.

The duct floor 2050 may also be adjusted by means of a hydraulic (orpneumatic) actuator 2095 comprised in the rotor 20. Hydraulic fluid canbe passed into the hydraulic actuator of the rotor 20 via the shaft 1280and a rotating type annular seal (not shown). In practice, the maximumrotational speed of a shaft relative to a rotational annular seal may belimited by a maximum rotational speed capability of the seal. Since, inembodiments, the rotor 20 is required to operate at a rotational speedsuch that the pinch point 50 is moving at a super sonic velocity, therotational speed of the shaft 1280 may be in excess of the rotationalspeed capability of the rotational seal. In such cases, an intermediaterotating member 2025 can optionally be placed axially adjacent to therotor 20 assembly (comprising the rotor 20 and at least the outputflange 1420), and operated at a rotational velocity which is below thatof the rotor 20. Hydraulic fluid can then be passed into theintermediate member 2025 by means of a rotational seal, and then fromthe intermediate member 2025 to the rotor 20 by means of a furtherrotational seal. Thus, by way of example when the intermediate member2025 is operated at a rotational speed equal to half that of the rotor20, each rotational seal need only be operated at a rotational speed ofhalf that which would otherwise be required if the intermediate member2025 was not employed. Of course, more than one intermediate member 2025can be used if necessary.

In a further embodiment, the duct floor 2050 can be actuatedmechanically, as shown in FIG. 22. For example, a wedge shaped cam 2210attached to a actuation shaft 2200 is arranged so that it can be movedaxially by axial movement of the actuation shaft. A cam follower 2220 isattached to the duct floor 2050. Thereby, in operation, by moving theactuation shaft 2200 axially, the wedge shaped cam 2210 is moved, andthe sloping surface of said wedge shaped cam 2210 acts on the camfollower 2220, thereby urging it in a direction perpendicular to the cam2210. The duct floor 2050 is thereby moved outwardly. A spring or othermeans of urging the duct floor 2050 towards the inward position may beemployed, so as to ensure that the cam follower 2220 remains in contactwith the cam 2210. Centrifugal force acting on the duct floor 2050 andcam follower 2220 may in some embodiments and operating conditionsovercome gas pressure in the duct 40, and thereby tend to force the ductfloor 2050 in a radially outward direction. The spring or equivalenturging member is arranged to resist this action. An alternativearrangement is that the cam follower 2220 is hook shaped 2225, orsimilar, such that it engages with a sloped surface 2215 of the cam 2210which is on the opposite side of the axis from the floor 2050. Thereby,the floor can be forced inwards by cam action. A further alternativearrangement is that the actuation shaft is rotated relative to the rotorshaft 2015 to which the rotor 20 is attached, and the actuation shaft2200 has an offset lobe shaped element (not shown) attached to its endinside the rotor 20, the lobe being similar to that employed on a camshaft of an internal combustion engine. Thus, in use, by rotation of theactuation shaft 2200 relative to the shaft 1280, in this embodiment thelobe is caused to rotate relative to the rotor 20, and the cam follower2220 or hook 2225 bears upon the lobe, thus forcing the floor 2050 in aradially outward or inward direction. Thus, an alternative arrangementfor adjusting the level of the duct floor 2050 is effected. Of course,other means of actuation such as electro-mechanical means can beemployed, using control coupling means equivalent to the rotatinghydraulic seals of the previously described hydraulic embodiment above.

It will be appreciated that any appropriate fluid, not limited to gas orliquid, can be operated on according to the invention, for purposes ofcompression, expansion or other operation on a property of the fluid.

Although the invention has been explained in relation to its preferredembodiments, these are not intended to limit the invention. It will beunderstood by those skilled in the art that many other modifications andvariations are possible without departing from the scope of theinvention as claimed. Embodiments and features of embodiments may bejuxtaposed or interchanged as appropriate.

1. A compressor or expander having a rotor and a stator forming a closedvolume therebetween and arranged for relative traversal, so as to form apinch point at the point of nearest proximity, wherein at least one ofthe rotor and the stator has a movable portion such that the closedvolume is controllably variable.
 2. An apparatus according to claim 1wherein the pinch point is arranged to move at substantially supersonicvelocity.
 3. An apparatus according to claim 1 wherein the rotor has avariable geometry formed by said movable portion.
 4. An apparatusaccording to claim 3 wherein said movable portion comprises a rotorsurface portion which is movable in a direction having a radialcomponent relative to an axis of rotation of the rotor.
 5. An apparatusaccording to claim 4 wherein the movable rotor surface portion isarranged to be actuated by fluid pressure.
 6. An apparatus according toclaim 5 wherein the fluid pressure is hydraulic pressure.
 7. Anapparatus according to claim 5 wherein the fluid pressure is pneumaticpressure.
 8. An apparatus according to claim 3 comprising a first flangearranged for rotation in the stator and for defining a boundary of theclosed volume.
 9. An apparatus according to claim 8 comprising a secondflange arranged for rotation in the stator and for defining anotherboundary of the closed volume.
 10. An apparatus according to claim 9wherein one of the flanges comprises said movable portion, movable in anaxial direction of the rotor.
 11. An apparatus according to claim 10wherein the movable flange is supported on an axially movable member.12. An apparatus according to claim 10 wherein the movable flange issupported in an axial groove in the rotor.
 13. An apparatus according toclaim 8 wherein the first flange has an inlet opening for allowing fluidto enter the compressor, the inlet opening being arranged such that inuse it trails the pinch point in the direction of rotation.
 14. Anapparatus according to claim 13 wherein a portion of the leading edge ofthe inlet opening extends at least partly away from the rotor in anaxial direction so as to form a scoop.
 15. An apparatus according toclaim 8 wherein the second flange has an outlet opening for allowingfluid to leave the compressor, the outlet opening being arranged suchthat in use it leads the pinch point in the direction of rotation. 16.An apparatus according to claim 1 wherein the rotor includes a labyrinthseal in the region of the pinch point.
 17. An apparatus according toclaim 15 comprising a third flange arranged for rotation in the statorand axially spaced from the second flange, so as to define an outputdiffusion chamber.
 18. An apparatus according to claim 17 wherein thestator comprises an outlet passage, and the output diffusion chamber andthe outlet passage are arranged for radial communication therebetween.19. An apparatus according to claim 17 wherein the output diffusionchamber comprises a ramp adjacent the outlet opening, the ramp orientedsuch that gas exiting the compressor axially through the outlet openingis redirected contrary to the direction of rotation.
 20. An apparatusaccording to claim 9 wherein each flange incorporates one or morelabyrinth seals.
 21. An apparatus according to claim 20 wherein thelabyrinth seal comprises circumferential grooves extending from a pointadjacent the pinch point to another point adjacent the pinch point andon the other side of the pinch point, such that the grooves are sealedin the region of the pinch point.
 22. An apparatus according to claim 1further comprising outlet valving means comprising a resilientdeformable disc arranged to seal against an axial end of the stator. 23.An apparatus according to claim 1 wherein the rotor is shaped so as toform a plurality of pinch points at respective points of proximity withthe stator.
 24. An apparatus according to claim 23 wherein the rotor isgenerally elliptical in shape.
 25. An apparatus according to claim 23wherein the generally elliptical rotor includes a relatively narrowcentral portion and lobes either side thereof.
 26. An apparatusaccording to claim 1 in which the rotor incorporates a passage arrangedto communicate axially with the housing, and with the duct.
 27. Anapparatus according to claim 1 wherein the rotor surface includes aportion adjacent to the pinch point in which proximity of the portion tothe stator varies as the rotor is traversed in an axial direction. 28.An apparatus according to claim 27 in which the proximity of the portiondecreases with distance along the axis from an inlet side.
 29. Anapparatus according to claim 27 in which the proximity of the portiondecreases with distance along the axis from an outlet side. 30-35.(canceled)
 36. A turbine or pump comprising a compressor or expanderaccording to claim
 1. 37. A method of compressing or expanding a fluidcomprising the steps of: arranging a rotor and a stator to form a closedvolume therebetween, and for relative traversal, so as to form a pinchpoint at the point of nearest proximity, and arranging at least one ofthe rotor and stator to have a movable portion, such that the closedvolume is controllably variable. 38-39. (canceled)